Method for the correction of the reduced mass flow rate of a compressor in an internal combustion engine turbocharged by means of a turbocharger

ABSTRACT

A method for correcting the reduced mass flow rate of a compressor in an internal combustion engine turbocharged by means of a turbocharger provided with a turbine and with a compressor; the internal combustion engine comprising an intake manifold and an exhaust manifold, and being set up to allow the passage of air from the intake manifold to the exhaust manifold; the method comprising determining, in a design stage, a control law that provides an objective opening of a control actuator of the wastegate as a function of an actual supercharging pressure and of a reduced mass flow rate of the compressor; and correcting the reduced mass flow rate of the compressor as a function of the enthalpy of a gas mixture flowing through the turbine of the turbocharger.

TECHNICAL FIELD

The present invention relates to a method for correcting the reducedmass flow rate of a compressor in an internal combustion engineturbocharged by means of a turbocharger.

PRIOR ART

As is known, some internal combustion engines are provided with aturbocharger supercharging system, which may increase the powerdeveloped by the engine by exploiting the enthalpy of exhaust gases tocompress the air aspirated by the engine and thus increase thevolumetric intake performance.

A turbocharger supercharging system comprises a turbocharger equippedwith a turbine, which is arranged along an exhaust pipe to rotate athigh speed under the propelling force of exhaust gases expelled from theengine, and with a compressor, which is driven in rotation by theturbine and is arranged along the air intake duct to compress the airaspirated by the engine. In a turbocharger supercharging system, it isnecessary to the keep the operating range of the turbocharger within auseful zone dependent on engine running conditions for both functionalreasons (i.e. to avoid irregular or, in any case, low efficiencyoperation) and structural reasons (i.e. to avoid damaging theturbocharger). In order to limit the supercharging pressure (i.e. thepressure of the compressed air downstream of the compressor), a bypassduct regulated by a wastegate (valve) is arranged in parallel with theturbine; when the wastegate opens, part of the exhaust gases flowthrough the bypass duct and thus bypass the turbine, this resulting in adrop in the rotor's rotational speed and a consequent decrease insupercharging.

The wastegate is operated by a pneumatic actuator, which is in turncontrolled by a solenoid control valve that enables regulating theaction of the wastegate. The pneumatic actuator comprises a sealedshell, which internally supports a flexible diaphragm that divides thesealed shell into two reciprocally fluid-tight chambers. The flexiblediaphragm is mechanically connected to a rigid rod that operates thewastegate to control the opening and closing of the wastegate. A firstchamber is connected to atmospheric pressure, while a second chamber isconnected to the supercharging pressure and may also be connected toatmospheric pressure through a duct controlled by the proportionalsolenoid control valve, which is able to choke the duct between a closedposition, in which the duct is completely closed, and a position ofmaximum opening.

A contrast spring is arranged in the first chamber such that it iscompressed between a wall of the shell and the flexible diaphragm 28 andrests against the flexible diaphragm on the side opposite to the rod.When the pressure difference between the two chambers is lower than anoperating threshold (determined by the contrast spring preload), the rodkeeps the wastegate in a completely closed position, while when thepressure difference between the two chambers is higher than theoperating threshold, the contrast spring starts to compress under thethrust of the flexible diaphragm, which thus deforms and causes adisplacement of the rod that consequently moves the wastegate towardsthe open position. By controlling the solenoid control valve, it ispossible to connect the second chamber to atmospheric pressure through avariable-sized opening, and so it is possible to adjust the pressuredifference between the two chambers that, in turn, causes the opening orclosing of the wastegate. It is important to note that until thedifference between the supercharging pressure and atmospheric pressureexceeds the operating threshold (equal to the preload generated by thecontrast spring divided by the area of the flexible diaphragm), thewastegate may not be opened by the action exerted by the solenoidcontrol valve (which may only reduce, and not increase, the differencebetween the supercharging pressure and atmospheric pressure).

In known internal combustion engines, an objective superchargingpressure is generated that is used to cause operation of the wastegateby adding an open-loop contribution and an closed-loop contribution: theopen-loop contribution is generated using an experimentally obtainedcontrol map, while the closed-loop contribution is provided by a PIDregulator that attempts to cancel a pressure error, i.e. a differencebetween the objective supercharging pressure and the actualsupercharging pressure measured by a sensor.

However, the preload generated by the contrast spring of the pneumaticactuator has high structure dispersion, considerable thermal drift andalso a certain time drift. Furthermore, the pneumatic actuator hasconsiderable hysteresis, i.e. the behaviour of the pneumatic actuatorvaries significantly between the opening movement and the oppositeclosing movement. In consequence, the control map used to determine theclosed-loop contribution is highly nonlinear and pursuing the objectivesupercharging pressure proves to be complicated; thus, pursuing theobjective supercharging pressure in known internal combustion enginestends to have large overshoots or undershoots (i.e. the actualsupercharging pressure significantly exceeds or drops below theobjective supercharging pressure) and thus causes oscillations,especially when the supercharging pressure is around the operatingthreshold, below which the wastegate may not be opened by the actionexerted by the solenoid control valve.

Overshoots (i.e. peaks) in the supercharging pressure are particularlytroublesome because they cause significant stress (and thereforepotential damage over time) to the mechanical components of the internalcombustion engine and because they generate both noise perceptible bythe vehicle occupants and corresponding undesired oscillations in thedrive torque generated by the internal combustion engine.

To reduce the size of the overshoots, it is possible to reduce theadditional contribution of the PID regulator used to calculate theclosed-loop contribution for operation of the wastegate.

For example, patent EP2314850 describes a method for controlling thewastegate comprising the steps of determining, in a design stage, acontrol law that provides an objective opening for a control actuator ofthe wastegate as a function of a supercharging pressure; determining anobjective supercharging pressure; measuring an actual superchargingpressure; determining a first open-loop contribution for an objectiveposition of the control actuator of the wastegate by means of thecontrol law and as a function of the objective supercharging pressure;determining a second closed-loop contribution for the objective positionof the control actuator of the wastegate; calculating the objectiveposition for the control actuator of the wastegate by adding the twocontributions; and controlling the control actuator of the wastegate soas to pursue the objective position for the control actuator of thewastegate.

The step of determining the second closed-loop contribution contemplatesdetermining a virtual position for control actuator of the wastegate bymeans of the control law and as a function of the actual superchargingpressure; calculating a position error by calculating the differencebetween the first open-loop contribution of the objective position forthe control actuator of the wastegate and the virtual position of thecontrol actuator of the wastegate; and determining the secondclosed-loop contribution by processing the position error by means of afirst regulator that attempts to cancel the position error.

However, the control method described in patent EP2314850 is quiterobust, quick and devoid of oscillations only in operating conditionswhere there is no significant passage of air directly from the intakemanifold to the exhaust of the internal combustion engine 1.

DESCRIPTION OF THE INVENTION

The object of the present invention is to provide a method forcorrecting the reduced mass flow rate of a compressor in an internalcombustion engine turbocharged by means of a turbocharger, thiscorrection method being devoid of the above-described drawbacks and, inparticular, simple and inexpensive to implement.

According to the present invention, a method for correcting the reducedmass flow rate of a compressor in an internal combustion engineturbocharged by means of a turbocharger is provided as claimed in theappended claims.

BRIEF DESCRIPTION OF DRAWINGS

The present invention will now be described with reference to theaccompanying drawings, which illustrate a non-limitative embodimentthereof, in which:

FIG. 1 is a schematic view of an internal combustion engine turbochargedby means of a turbocharger and equipped with a control unit thatimplements the method for correcting the reduced mass flow rate of acompressor, the subject of the present invention;

FIG. 2 is a schematic view of a pneumatic actuator of the wastegate;

FIG. 3 is a graph showing an experimental control map; and

FIG. 4 is a block diagram of a control logic of the wastegate.

PREFERRED EMBODIMENTS OF THE INVENTION

In FIG. 1, reference numeral 1 indicates, as a whole, an internalcombustion engine supercharged by means of a turbocharger superchargingsystem 2.

The internal combustion engine 1 comprises four cylinders 3, each ofwhich is connected to an intake manifold 4 by at least one respectiveinlet valve (not shown) and to an exhaust manifold 5 by at least onerespective exhaust valve (not shown). The intake manifold 4 receivesfresh air (i.e. air coming from the external environment) through anintake duct 6, which is fitted with an air filter 7 and is regulated bya butterfly valve 8. An intercooler 9 for cooling the aspirated air isarranged along the intake duct 6. An exhaust pipe 10 is connected to theexhaust manifold 5 to feed the exhaust gases produced by combustion toan exhaust system, which emits the gases produced by combustion into theatmosphere and normally comprises at least one silencer (not shown)arranged downstream of the catalytic converter 11.

The supercharging system 2 of the internal combustion engine 1 comprisesa turbocharger 12 equipped with a turbine 13, which is arranged alongthe exhaust pipe 10 in order to rotate at high speed under the action ofthe exhaust gases expelled from the cylinders 3, and a compressor 14,which is arranged along the intake duct 6 and is mechanically connectedto the turbine 13 in order to be drawn in rotation by the turbine 13 andthus increase the pressure of the air fed through the intake duct 6.

A bypass duct 15 is provided along the exhaust pipe 10 and is connectedin parallel to the turbine 13 such that its ends are connected upstreamand downstream of the turbine 13; a wastegate 16, suitable forregulating the flow of the exhaust gases running through the bypass duct15 and controlled by a pneumatic actuator 17, is arranged along thebypass duct 15. A bypass duct 18 is provided along the intake duct 6 andis connected in parallel to the compressor 14 such that its ends areconnected upstream and downstream of the compressor 14; a Poff valve 19,suitable for regulating the flow of air running through the bypass duct18 and controlled by an electric actuator 20, is arranged along thebypass duct 18.

The internal combustion engine 1 is controlled by an electronic controlunit 21, which superintends the operation of all the components of theinternal combustion engine 1, including the supercharging system 2. Inparticular, the electronic control unit 21 controls the actuators 17 and20 of the wastegate 16 and of the Poff valve 19. The electronic controlunit 21 is connected to sensors 22 that measure the temperature andpressure along the intake duct 6 upstream of the compressor 14, tosensors 23 that measure the temperature and pressure along the intakeduct 6 upstream of the butterfly valve 8, and to sensors 24 that measurethe temperature and pressure inside the intake manifold 4. In addition,the electronic control unit 21 is connected to a sensor 25 that measuresthe angular position (and thus the speed of rotation) of a crankshaft ofthe internal combustion engine 1 and a sensor 26 that measures thetiming of the intake and/or exhaust valves.

As shown in FIG. 2, the pneumatic actuator 17 of the wastegate 16comprises a sealed shell 27 that internally supports a flexiblediaphragm 28, which divides the sealed shell 27 into two reciprocallyisolated chambers 29 and 30. The flexible diaphragm 28 is mechanicallyconnected to a rigid rod 31 that operates the wastegate 16 to controlthe opening and closing of the wastegate 16. Chamber 29 is connected bymeans of a duct 32 to atmospheric pressure (taken upstream of thecompressor 14), while chamber 30 is connected by means of a duct 33 tothe supercharging pressure (taken downstream of the compressor 14) andis connected by means of a duct 34 to atmospheric pressure (takenupstream of the compressor 14). Duct 34 is not clear, but is regulatedby a solenoid control valve 35 that may choke the duct 34 between aclosed position, in which the duct 34 is completely closed, and amaximum opening position.

A contrast spring 36 is arranged in chamber 29 such that it iscompressed between a wall of the shell 27 and the flexible diaphragm 28and rests against the flexible diaphragm 28 on the side opposite to therod 31. When the pressure difference between chamber 30 and chamber 29is lower than an operating threshold (determined by the preload of thecontrast spring 36), the rod 31 keeps the wastegate 16 in a completelyclosed position, while when the pressure difference between chamber 30and chamber 29 is higher than the operating threshold, the contrastspring 36 starts to compress under the thrust of the flexible diaphragm28, which thus deforms and causes a displacement of the rod 31 thatconsequently moves the wastegate 16 towards the open position. Bycontrolling the solenoid control valve 35, it is possible to connectchamber 30 to atmospheric pressure through a variable-sized opening, andso it is possible to adjust the pressure difference between the twochambers 29 and 30 that, in turn, causes the opening or closing of thewastegate 16.

It is important to note that until the difference between thesupercharging pressure P and atmospheric pressure P_(atm) exceeds theoperating threshold (equal to the preload generated by the contrastspring 36 divided by the area of the flexible diaphragm 28), thewastegate 16 may not be opened by the action exerted by the solenoidcontrol valve 35 ((which may only reduce, and not increase, thedifference between the supercharging pressure P and atmospheric pressureP_(atm)). Due to structure dispersion, thermal drift and time drift, thepreload generated by the contrast spring 36 is only known with quite ahigh level of uncertainty (in the order of ±20%).

During a design stage of the internal combustion engine 1, a control lawCL is experimentally determined that provides an objective opening WGfor the solenoid control valve 35 of the wastegate 16 as a function of asupercharging pressure P (or rather a supercharging ratio RP that isequal to the ratio between the supercharging pressure P and atmosphericpressure P_(atm) and is equivalent to the supercharging pressure P) anda reduced mass flow rate M_(R) of the compressor 14. In other words, thecontrol law CL provides the objective opening WG for the solenoidcontrol valve 35 of the wastegate 16 that should allow achieving adesired supercharging pressure P (or rather a desired superchargingratio RP) in the presence of a given reduced mass flow rate M_(R).According to a preferred embodiment shown by way of example in FIG. 3,the control law CL consists of an experimental map (i.e. a table orrather a matrix, which is highly linear, as is evident in FIG. 3);alternatively, the control law CL may consist of an arithmeticalfunction. The control law CL is stored in a memory of the electroniccontrol unit 21 for subsequent use as described below.

In use, during normal operation of the internal combustion engine 1, theelectronic control unit 21 measures the actual supercharging pressure P(i.e. the air pressure along the intake duct 6 downstream of thecompressor 14), measures or estimates (in a known manner) theatmospheric pressure P_(atm), and estimates (in a known manner) theactual reduced mass flow rate M_(R) of the compressor 14. Furthermore,during normal operation of the internal combustion engine 1, theelectronic control unit 21 determines, in a known manner, an objectivesupercharging pressure P_(obj), which must be pursued by controlling, ifneeded, the solenoid control valve 35 of the wastegate 16. In order tocontrol the solenoid control valve of the wastegate 16, the electroniccontrol unit 21 determines an objective position WG_(obj) for thesolenoid control valve 35 of the wastegate 16, which is normallyactuated using an open-loop control.

As shown in FIG. 4, the objective position WG_(obj) for the solenoidcontrol valve 35 of the wastegate 16 is calculated by algebraicallyadding (i.e. taking the sign into account) four contributions: anopen-loop contribution WG_(OL), a closed-loop contribution WG_(CL1), aclosed-loop contribution WG_(CL2), and an adaptive contribution WG_(A).

The open-loop contribution WG_(OL) is determined using the control lawCL: an objective compression ratio RP_(obj) (equal to the ratio betweenthe objective supercharging pressure P_(obj) and atmospheric pressureP_(atm) and equivalent to the objective supercharging pressure P_(obj))is determined as a function of the objective supercharging pressureP_(obj); the objective compression ratio RP_(obj) and the actual reducedmass flow rate M_(R) are then supplied to computation block 37 that, byusing the control law CL, provides the open-loop contribution WG_(OL).

Preferably, before being supplied to computation block 37, the objectivecompression ratio RP_(obj) is filtered by means of a first-orderlow-pass filter 38 to reduce the speed of change; in other words, theobjective compression ratio RP_(obj) is filtered by means of low-passfilter 38 so as to slow down the evolution of the objective compressionratio RP_(obj), in this way “rounding off” any step changes. Thefunction of the low-pass filter 38 is to make the evolution of theobjective compression ratio RP_(obj) more “real” (i.e. closer to whathappens in reality), as it is clear that step changes (or in any casevery rapid ones) in the actual supercharging pressure P are not possiblebecause of the obvious physical limits due to the inertia involved.According to a preferred embodiment, a cutoff frequency of the low-passfilter 38 is determined as a function of the reduced mass flow rateM_(R) of the compressor and the actual supercharging ratio RP accordingto an experimentally determined law.

According to a preferred embodiment, the open-loop contribution WG_(OL)supplied by computation block 37 is first compensated by means of threecompensation parameters K_(atm), K_(H2O) and K_(air) and then filteredby means of a first-order low-pass filter 39 to reduce speed of change.Compensation parameters K_(air) is determined by computation block 40 asa function of the temperature T_(air) of the aspirated air and using alinear equation having experimentally determined coefficients,compensation parameter K_(H2O) is determined by computation block 41 asa function of the temperature T_(H2O) of a cooling liquid of theinternal combustion engine 1 and using a linear equation havingexperimentally determined coefficients, and compensation parameterK_(atm) is determined by computation block as a function of atmosphericpressure P_(atm) and using a linear equation having experimentallydetermined coefficients; the coefficients of the linear equation thatprovides compensation parameter K_(atm) as a function of atmosphericpressure P_(atm) might not be constant, but vary as a function of thereduced mass flow rate M_(R) of the compressor 14 and the actualsupercharging ratio RP according to an experimentally determined law.

The open-loop contribution WG_(OL) is filtered by means of the low-passfilter 39 so as to slow down the evolution of the open-loop contributionWG_(OL), in this way “rounding off” any step changes. The function ofthe low-pass filter 39 is to make the evolution of the open-loopcontribution WG_(OL) more “real” (i.e. closer to what happens inreality), as it is clear that step changes (or in any case very rapidones) in the position of the solenoid control valve 35 are not possiblebecause of the obvious physical limits due to the inertia involved.According to a preferred embodiment, a cutoff frequency for low-passfilter 39 is determined as a function of the actual supercharging ratioRP according to an experimentally determined law. According to apreferred embodiment, the open-loop contribution WG_(OL) isasymmetrically filtered by means of the low-pass filter 39: theopen-loop contribution WG_(OL) is filtered by means of the low-passfilter 39 only when the open-loop contribution WG_(OL) varies to openthe wastegate 16 and not when the open-loop contribution WG_(OL) variesto close the wastegate 16; in this manner, the action of the compressor14 is more rapid (more reactive), favouring the responsiveness of theinternal combustion engine (thereby reducing turbo-lag), while thestopping of the compressor 14 is softer. It should be noted that whenmaximum performance is sought, an “abrupt” reaction of the internalcombustion engine 1 is acceptable (and in some cases even desired),while in other cases “softer” behaviour, i.e. without excessively rapidand sharp reactions, is desired. It should also be noted that by virtueof the presence of the low-pass filter 39, possible oscillatoryphenomena in the pneumatic actuator 17 of the wastegate 16 are eithereliminated or greatly attenuated; this result is achieved by virtue ofthe fact that the action of the low-pass filter 39 avoids passingexcessively rapid stress variations, which could trigger oscillatoryphenomena, to the flexible diaphragm 28 and the contrast spring 36.

The closed-loop contribution WG_(CL1) of the objective position WG_(obj)of the solenoid control valve 35 of the wastegate 16 is obtained byusing a virtual position WGF of the wastegate 16 (therefore a controlquantity that has no precise correspondence with physical reality) as afeedback variable, which is not determined by direct measurement with areal measurement sensor, but by using the control law CL as ameasurement sensor. In other words, a computation block 43 supplies thevirtual position WGF of the wastegate 16 by applying the control law CLbased on the actual supercharging pressure P (or rather the actualsupercharging ratio RP) and the reduced mass flow rate M_(R) of thecompressor 14; therefore, the virtual position WGF of the wastegate 16corresponds to the position that the wastegate 16 should have accordingto the control law CL (and thus affected by all the errors of thecontrol law CL) in conjunction with the actual supercharging ratio RPand the actual reduced mass flow rate M_(R) of the compressor 14. Thevirtual position WGF of the wastegate 16 is compared with the open-loopcontribution WG_(OL), which corresponds to the position that thewastegate 16 should have according to the control law CL (and thusaffected by all the errors of the control law CL) in conjunction withthe objective compression ratio RP_(obj) and the actual reduced massflow rate M_(R) of the compressor 14; in other words, the open-loopcontribution WG_(OL) represents an objective for the virtual positionWGF as it is calculated using the objective compression ratio RP_(obj).In particular, a position error ε_(WG) is computed by calculating thedifference between the open-loop contribution WG_(OL) of the objectiveposition WG_(obj) of the solenoid control valve 35 of the wastegate 16and the virtual position WGF of the wastegate 16, and this positionerror ε_(WG) is supplied to a PID regulator 44, which attempts to cancelthe position error ε_(WG).

The fact of comparing two values (the open-loop contribution WG_(OL),which represents an objective for the virtual position WGF, and thevirtual position WGF) obtained from the control law CL allows tocompensate the errors of the control law CL and to linearize the highlynonlinear behaviour of the wastegate 16; in this manner, the PIDregulator 44 may work more stably and the calibration of the controlparameters (i.e. of the proportional, integrative and derivativecoefficients and the saturation thresholds) of the PID regulator 44 isrelatively simple. Furthermore, the control loop of the PID regulator 44is self-compensated with respect to the temperature T_(air) of theaspirated air, the temperature T_(H2O) of a cooling liquid of theinternal combustion engine 1, and atmospheric pressure P_(atm).

The closed-loop contribution WG_(CL2) of the objective position WG_(obj)of the solenoid control valve 35 of the wastegate 16 is determined byusing the supercharging pressure P as a feedback variable; a pressureerror ε_(p) is then computed by calculating the difference between theobjective supercharging pressure P_(obj) and the actual superchargingpressure P, and the pressure error ε_(p) is supplied to a PID regulator45, which attempts to cancel the pressure error ε_(p).

Preferably, before being compared with the actual supercharging pressureP, the objective supercharging pressure P_(obj) is filtered by means ofa first-order low-pass filter 46 to reduce the speed of change; in otherwords, the objective supercharging pressure P_(obj) is filtered by meansof low-pass filter 46 so as to slow down the evolution of the objectivesupercharging pressure P_(obj), in this way “rounding off” any stepchanges. The function of the low-pass filter 46 is to make the evolutionof the objective supercharging pressure P_(obj) more “real” (i.e. closerto what happens in reality), as it is clear that step changes (or in anycase very rapid ones) in the actual supercharging pressure P are notpossible because of the obvious physical limits due to the inertiainvolved. According to a preferred embodiment, a cutoff frequency forlow-pass filter 46 is determined as a function of the reduced mass flowrate M_(R) of the compressor 14 and the actual supercharging ratio RPaccording to an experimentally determined law.

To avoid negative interference between the action of PID regulator 44and the action of PID regulator 45, the dynamics of PID regulator 44 isdifferent from the dynamics of PID regulator 45; in particular, PIDregulator 44 is essentially proportional and derivative (i.e. has highproportional and derivative coefficients and a low integral coefficient)in order to be ready (i.e. to work rapidly), while PID regulator isessentially integral (i.e. has low proportional and derivativecoefficients and a high integral coefficient) in order to guaranteeconvergence between the objective supercharging pressure P_(obj) and theactual supercharging pressure P. Thus, PID regulator 44 is used to reactrapidly and promptly to variation in the objective superchargingpressure P_(obj), while PID regulator 45 is used make the actualsupercharging pressure P converge with the objective superchargingpressure P_(obj) at the end of the transient.

The adaptive contribution WG_(A) of the objective position WG_(obj) ofthe solenoid control valve 35 of the wastegate 16 is essentially a“historical memory” of previous operation of the wastegate 16 and takespast control actions into account. The adaptive contribution WG_(A) isstored in a memory 47 of the electronic control unit 21 and iscyclically updated when the turbocharger 12 is at a steady speed (forexample, when the reduced mass flow rate M_(R) of the compressor 14 andthe supercharging ratio RP remain approximately constant for at least apredetermined interval of time) by using an integral term of PIDregulator 45 and/or PID regulator 44; in substance, the adaptivecontribution WG_(A) is equal to an “average” of past integral terms ofPID regulator 45 and/or PID regulator 44 when the turbocharger 12 wasunder steady speed conditions. When the turbocharger 12 is in steadyrunning conditions, the adaptive contribution WG_(A) stored in thememory 47 is updated by using the integral term of PID regulator 45and/or PID regulator 44 weighted by a weight W that is essentially basedon an actual position WG of the solenoid control valve 35 of thewastegate 16 in such a way that the weight W is minimum when thehysteresis in controlling the wastegate 16 is maximum; in this manner,adaptation is always gradual (i.e. the last integral term of PIDregulator 45 and/or PID regulator 44 may not distort the adaptivecontribution WG_(A) stored in the memory 47) and loadinghysteresis-distorted values in the adaptive contribution WG_(A) isavoided.

Generally, the adaptive contribution WG_(A) varies as a function of thereduced mass flow rate M_(R) of the compressor 14 and the superchargingratio RP. Furthermore, the adaptive contribution WG_(A) is filtered bymeans of a first-order low-pass filter 48 to reduce the speed of change;in other words, the adaptive contribution WG_(A) is not suppliedabruptly, but is supplied gradually to avoid step-like operation, whichnever corresponds to the physical reality, and thus to favour controlconvergence. According to a preferred embodiment, the cutoff frequencyof low-pass filter 48 is constant; alternatively, the cutoff frequencyof low-pass filter 48 could be altered as a function of the reduced massflow rate M_(R) of the compressor 14 and the supercharging ratio RP.

The integral term of the PID regulators 44 and 45 incorporates a“memory” of the errors that occurred in the immediate past; therefore,when variations in the surrounding conditions occur, the “memory” of theerrors that occurred in the immediate past contained in the integralterm of PID regulators 44 and 45 may have negative effects because itrepresents a situation that is no longer present.

The electronic control unit 21 resets (or possibly “freezes”, i.e.prevents a further growth of) each integral term of the PID regulators44 and 45 in the case of rapid change, i.e. a high transient, in theobjective supercharging pressure P_(obj) if the integral term itself ishigh, i.e. higher than the absolute value of a predetermined threshold;in other words, when a rapid change occurs in the objectivesupercharging pressure P_(obj) and an integral term of the PIDregulators 44 and 45 is higher in absolute value than a predeterminedthreshold, then the integral term is either reset or frozen (i.e. it isnot changed until the end of the high transient).

To establish whether a high transient of the objective superchargingpressure P_(obj) is present (i.e. a rapid change in the objectivesupercharging pressure P_(Obj)) the electronic control unit 21 comparesthe objective supercharging pressure P_(obj) with an objectivesupercharging pressure P_(obj-F) filtered by low-pass filter 49 todetermine a gradient ΔP_(obj) of the objective supercharging pressureP_(obj) that indicates the speed of change of the objectivesupercharging pressure P_(obj). In other words, gradient ΔP_(obj) of theobjective supercharging pressure P_(obj) is computed by calculating thedifference between the objective supercharging pressure P_(obj) and theobjective supercharging pressure P_(obj) filtered by low-pass filter 49.When gradient ΔP_(obj) of the objective supercharging pressure P_(obj)is higher than a threshold value, the electronic control unit 21 thenestablishes the presence of a high transient of the objectivesupercharging pressure P_(obj) (i.e. of a rapid change in the objectivesupercharging pressure P_(obj)) and thus resets (or possibly “freezes”)the integral terms of the PID regulators 44 and 45; this threshold valuemay be based on the supercharging ratio RP and the reduced mass flowrate M_(R) of the compressor 14. According to a preferred embodiment, acutoff frequency of the low-pass filter 49 is determined as a functionof the reduced mass flow rate M_(R) of the compressor 14 and the actualsupercharging ratio RP according to an experimentally determined law.

According to a preferred embodiment, the electronic control unit 21alters the integral coefficients of the PID regulators 44 and 45 as afunction of the pressure error ε_(p), so as alter the controlcharacteristics as the magnitude of the pressure error ε_(p) varies. Inparticular, the electronic control unit 21 alters the integralcoefficients of the PID regulators 44 and 45 in a manner inverselyproportional to the pressure error ε_(p), such that the smaller thepressure error ε_(p) the higher the integral coefficients of the PIDregulators 44 and 45, and alters the proportional coefficients of thePID regulators 44 and 45 in a manner directly proportional to thepressure error ε_(p), such that the larger the pressure error ε_(p) thehigher the proportional coefficients of the PID regulators 44 and 45. Inother words, the integral term of the PID regulators 44 and 45 (directlyproportional to the integral coefficients of the PID regulators 44 and45) serves to guarantee convergence between the actual superchargingpressure P and the objective supercharging pressure P_(obj), but thisconvergence is reached at the end of a transitory when the pressureerror ε_(p) is relatively small; at the beginning of the transient whenthe pressure error ε_(p) is large, the integral term of the PIDregulators 44 and 45 may generate oscillations and so, to avoid thisrisk, the integral coefficients of the PID regulators 44 and 45 arereduced at the beginning of the transient when the pressure error ε_(p)is large. The opposite applies to the proportional terms of the PIDregulators 44 and 45 (directly proportional to the proportional andderivative coefficients of the PID regulators 44 and 45), which must behigh when the pressure error ε_(p) is large to ensure rapid response andmust be low when the pressure error ε_(p) is small to ensureconvergence.

In the above-described low-pass filters 38, 46 and 49, the cutofffrequency is determined as a function of the reduced mass flow rateM_(R) of the compressor 14 and the actual supercharging ratio RP;according to an equivalent embodiment, the cutoff frequency isdetermined as a function of the speed of rotation of the internalcombustion engine 1 and of a gear engaged in a transmission driven bythe internal combustion engine 1. In this regard, it is important tonote that the dynamics of the turbocharger 12 varies significantlyaccording to the gear engaged, as the rise in the speed of rotation ofthe internal combustion engine 1 is rapid in the low gears, and so theincrease in rotational speed of the turbocharger 12 is equally rapid;instead, in the high gears, the rise in the speed of rotation of theinternal combustion engine 1 is slow, and so the increase in rotationalspeed of the turbocharger 12 is equally slow.

Similarly, the threshold value with which the gradient ΔP_(obj) of theobjective supercharging pressure P_(obj) is compared to establishwhether a high transient of the objective supercharging pressure P_(obj)is present may also be a function of the reduced mass flow rate M_(R) ofthe compressor 14 and the actual supercharging ratio RP or it could be afunction of the speed of rotation of the internal combustion engine 1and a gear engaged in a transmission driven by the internal combustionengine 1.

It should be pointed out that the supercharging pressure P and thesupercharging ratio RP are perfectly equivalent to one another, becausethe atmospheric pressure P_(atm) is approximately constant and has avalue approximating to a unitary value; therefore, using thesupercharging ratio RP is equivalent to using the supercharging pressureP and vice versa. In the control chart shown in FIG. 4 and describedabove, the supercharging ratio RP is used, but according to anequivalent embodiment (not shown), it is possible to use thesupercharging pressure P instead of the supercharging ratio RP.

In the embodiment described above, the control law CL provides anobjective opening WG for the solenoid control valve 35 of the wastegate16 as a function of a supercharging pressure P (or rather asupercharging ratio RP that is equal to the ratio between thesupercharging pressure P and atmospheric pressure P_(atm) and isequivalent to the supercharging pressure P) and a reduced mass flow rateM_(R) of the compressor 14; according to an equivalent embodiment, thecontrol law CL provides an objective opening WG for the solenoid controlvalve 35 of the wastegate 16 as a function of the power delivered by theinternal combustion engine 1 and volumetric efficiency of the internalcombustion engine 1, or as a function of a speed of rotation of theinternal combustion engine 1 and volumetric efficiency of the internalcombustion engine 1 (obviously different combinations of the parametersof the internal combustion engine 1 are also possible).

The electronic control unit 21 is also set up to control theturbocharged internal combustion engine 1 in order to reduce as much aspossible the phenomenon known as turbolag, i.e. the turbochargingresponse delay of the turbocharger 12.

In particular, the electronic control unit 21 is set up to control theturbocharged internal combustion engine 1 to operate in scavenging mode,in which a significant passage of air directly from the intake manifold4 to the exhaust pipe 10 of the internal combustion engine 1 iscontemplated.

Two possible operating configurations are thus possible, of which afirst configuration hereinafter indicated as the normal or traditionalconfiguration and a second configuration hereinafter indicated as thescavenging configuration, in which there is significant passage of airdirectly from the intake manifold 4 to the exhaust pipe 10 of theinternal combustion engine 1.

Typically, the passage of air directly from the intake manifold 4 to theexhaust pipe 10 is implemented by means of opportune timing of the inletvalves (not shown) that connect each cylinder 3 to the intake manifold 4and the exhaust valves (not shown) that connect each cylinder 3 to theexhaust manifold 5 to allow the passage of fresh air directly from theintake manifold 4 to the exhaust manifold 5 and then to the exhaust pipe10 of the internal combustion engine 1.

It is evident that the operation of the inlet valves (not shown) thatconnect each cylinder 3 to the intake manifold 4 and the exhaust valves(not shown) that connect each cylinder 3 to the exhaust manifold 5 maybe implemented by means of a known type of actuator, such as, forexample, a VVT (Variable Valve Timing) actuator, a camlesselectromagnetic actuator, or an electrohydraulic actuator.

According to s further variant, the electronic control unit 21 is alsoset up to control the turbocharged internal combustion engine 1 in orderto increase the mass and volumetric flow rates of air and/or exhaustgases that flow through the compressor 14 and the turbine 13, withrespect to the flow rate of air actually used by the turbochargedinternal combustion engine 1 in combustion to generate the desired powerlevel.

In order to implement the aforesaid control strategy, the electroniccontrol unit 21 is set up to differentiate the running of the cylinders3, in particular, to differentiate the flow rate of aspirated air andair trapped by each cylinder 3, and to differentiate the operating mode.The electronic control unit 21 is set up to generate the objectivetorque required by the driver of the vehicle with only some of thecylinders 3 firing, while the remaining cylinders 3 suck in as much airas possible. For example, in a turbocharged internal combustion engine 1with four cylinders 3, two cylinders 3 are active and produce thedesired torque by sucking in a mass of air that is approximately doublewith respect to the mass of air they would suck in under normal runningconditions (namely in the case where all four cylinders 3 are active).The two remaining cylinders 3 are not active and are controlled to suckin the most air, but are not involved in combustion. The mass of airthat flows through the two inactive cylinders does not take part incombustion and passes directly from the intake manifold to the exhaust.

Two possible operating configurations are thus possible, of which afirst configuration with four actively firing cylinders (hereinafterindicated as the normal configuration) and a second configuration withtwo cylinders 3 actively firing and two cylinders 3 that are controlledin air intake, but are not involved in fuel injection or combustion(hereinafter indicated as the virtual scavenging configuration).

The control strategies of the scavenging or virtual scavengingconfiguration and the normal configuration are described in patentapplications BO2012A000322, BO2012A000323 and BO2012A000324, entirelyincorporated herein for reference.

In the case where the internal combustion engine 1 is in the scavengingor virtual scavenging configuration, the following relation holds:

m=m _(com) +m _(scav)  [1]

m _(scav) =m−m _(com)  [1*]

where,m: total air mass flow rate (preferably low) through the internalcombustion engine 1;m_(com): mass flow rate of air trapped in the active cylinders 3 thattake part in combustion; andm_(scav): mass flow rate of air that is not involved in combustion andflows directly from the intake manifold 4 to the exhaust. Let us nowconsider a generic differential formula for the enthalpy at constantpressure per unit mass, according to which:

dh(T)=Cp*dT  [2]

where,h: enthalpy per unit mass;Cp: specific heat of gas at a constant pressure; andT: gas temperature.

By using formula [2], it is possible to obtain the differential formulafor the enthalpy of the total air mass flow rate m (preferably low)through the internal combustion engine 1, according to which:

dH(T)=m*Cp*dT  [3]

where,H: enthalpy of total air mass flow rate m (preferably low) through theinternal combustion engine 1;m: total air mass flow rate (preferably low) through the internalcombustion engine 1;Cp: specific heat of gas at a constant pressure; andT: gas temperature.

Integrating formula [3] gives the following:

H(T)=m*Cp*(T−T _(ref))  [4]

where,H: enthalpy of total air mass flow rate m (preferably low) through theinternal combustion engine 1;m: total air mass flow rate (preferably low) through the internalcombustion engine 1;Cp: specific heat of gas at a constant pressure;T: gas temperature; andT_(ref): reference temperature of gas.

Assuming that the temperature T_(ref) of the gas is equal to zero, it isthen possible to simplify formula [4] to give:

H(T)=m*Cp*T  [5]

where,H: enthalpy of total air mass flow rate m (preferably low) through theinternal combustion engine 1;m: total air mass flow rate (preferably low) through the internalcombustion engine 1;Cp: specific heat of gas at a constant pressure; andT: gas temperature.

In the case where the internal combustion engine 1 is in the scavengingconfiguration, the enthalpy value of the exhaust gases is obtained bysubstituting terms in formula [5] to give the following formula:

H _(com)(T _(ext))−M _(com) *Cp*T _(ext)  [6]

where,H_(com): enthalpy of mass flow rate m_(com) of air trapped in thecylinders 3 that participates in combustion;m_(com): mass flow rate of air trapped in the cylinders 3 thatparticipates in combustion;Cp: specific heat of gas at a constant pressure; andT_(ext): temperature of exhaust gases reaching the exhaust system andestimated by the electronic control unit.

The enthalpy value of the mixture of exhaust gases and fresh air isobtained by substituting terms in formula [5] to give the followingformula:

H(T _(mixt))=m*Cp*T _(mixt)  [7]

where,H: enthalpy of total air mass flow rate m (preferably low) through theinternal combustion engine 1;m: total air mass flow rate (preferably low) through the internalcombustion engine 1;Cp: specific heat of gas at a constant pressure; andT_(mixt): temperature of the mixture of exhaust gases and fresh air.

The temperature T_(mixt) of the mixture of exhaust gases and fresh airmay be calculated with the following formula:

T _(mixt)=(T _(ext) *m _(com) +T _(air) *m _(scav))m  [8]

where,T_(mixt): temperature of the mixture of exhaust gases and fresh air;m: total air mass flow rate (preferably low) through the internalcombustion engine 1;T_(ext): temperature of exhaust gases reaching the exhaust system andestimated by the electronic control unit;m_(com): mass flow rate of air trapped in the cylinders 3 thatparticipates in combustion;T_(air): temperature of air inside the intake manifold 4 measured by atemperature sensor arranged inside the intake manifold 4; andm_(scav): mass flow rate of air that is not involved in combustion andflows directly from the intake manifold 4 to the exhaust.

The temperature T_(mixt) of the mixture of exhaust gases and fresh airmay be calculated with formula [7], which represents the mixing balanceof two gas masses that have different temperatures, although it ispossible to assume that they have the same pressure and the samespecific heat Cp at constant pressure.

From the ratio between the enthalpy value of the mixture of exhaustgases and fresh air obtained with formula [7] and the enthalpy value ofthe exhaust gases obtained with formula [6], it follows that:

H(T _(mixt))H _(com)(T _(ext))=(m−*Cp*T _(mixt))/(m _(com) *Cp*T_(ext))  [9]

Substituting the temperature T_(mixt) of the mixture of exhaust gasesand fresh air calculated with formula [8] in the just obtained equation[9], gives:

$\begin{matrix}\begin{matrix}{{{H\left( T_{mixt} \right)}/{H_{com}\left( T_{ext} \right)}} = {\left( {{m*T_{ext}*m_{com}} + {m*T_{air}*m_{scav}}} \right)/}} \\{\left( {m*m_{com}*T_{ext}} \right)} \\{= {\left( {{T_{ext}*m_{com}} + {T_{air}*m_{scav}}} \right)/\left( {m_{com}*T_{ext}} \right)}} \\{= {1 + {\left( {T_{air}*m_{scav}} \right)/\left( {m_{com}*T_{ext}} \right)}}} \\{= {1 + {\left( {T_{air}/T_{ext}} \right)*\left( {m_{scav}/m_{com}} \right)}}}\end{matrix} & \lbrack 10\rbrack\end{matrix}$

Substituting, by means of formula [1*], the mass flow rate m_(scav) ofair that is not involved in combustion and flows directly from theintake manifold 4 to the exhaust in the just obtained equation [10]gives:

=1+(T _(air) /T _(ext))*((m−m _(com))/m _(com))

=1+(T _(air) /T _(ext))*(m/m _(com)−1)

=1+(T _(air) /T _(ext))*(η_(scav)−1)  [11]

Where η_(scav) represents the scavenging efficiency of the cylinders 3and is obtained from the ratio between the total air mass flow rate m(preferably low) through the internal combustion engine 1 and the massflow rate m_(com) of air trapped inside the cylinders 3 thatparticipates in combustion.

Finally, from formula [11], it is possible to calculate the enthalpyvalue of the mixture of exhaust gases and fresh air as follows:

H(T _(mixt))=H _(com)(T _(ext))*[1+(T _(air) /T_(ext))*(η_(scav)−1)]  [12]

A corrective enthalpy value H_(corr) _(—) ₁ may then be defined that isequal to:

H _(corr) _(—) ₁=[1+(T _(air) /T _(ext))*(η_(scav)−1)]  [13]

It is important to stress that the hitherto described method fordetermining the corrective enthalpy value H_(corr) _(—) ₁ contemplatesassuming the following working conditions:

-   -   the exhaust pressure remains substantially constant;    -   ideal gases are used; and    -   the fuel mass m_(fuel) is negligible.

The corrective enthalpy value H_(corr) _(—) ₁ is used in the control lawCL, which provides the objective opening WG for the solenoid controlvalve 35 of the wastegate 16 as a function of the supercharging pressureP (or rather a supercharging ratio RP that is equal to the ratio betweenthe supercharging pressure P and atmospheric pressure P_(atm) and isequivalent to the supercharging pressure P) and the reduced mass flowrate M_(R) of the compressor 14. In particular, the corrective enthalpyvalue H_(corr) _(—) ₁ is used in the control law CL for correcting thereduced mass flow rate M_(R) of the compressor 14; it is advisable touse the corrective enthalpy value H_(corr) _(—) ₁ for correcting thereduced mass flow rate M_(R) of the compressor 14 because the expansionwork in the turbine 13 is substantially equal to the enthalpy jump ofthe fluid flowing through the turbine 13.

The correct reduced mass flow rate M_(R) _(—) _(corr) may be calculatedas follows:

M _(R) _(—) _(corr) =M _(R)/η_(scav) *H _(corr) _(—) ₁  [14]

where,M_(R) _(—) _(corr): correct reduced mass flow rate of the compressor 14;M_(R): reduced mass flow rate of the compressor 14; andH_(corr) _(—) ₁: first corrective enthalpy value calculated with formula[13].

It is therefore possible to correct the reduced mass flow rate M_(R) ofthe compressor 14 as a function of the enthalpy of the gas mixtureflowing through the turbine 13 of the turbocharger 12.

It is important to stress that the reduced mass flow rate M_(R) of thecompressor 14 corresponds to the total air mass flow rate m (preferablylow) through the internal combustion engine 1 used in the foregoingdescription regarding formulae [1] to [13].

According to a further variant, the correct reduced mass flow rate M_(R)_(—) _(corr) of the compressor 14 may be calculated as follows:

M _(R) _(—) _(corr) =M _(R)/η_(scav) *H _(corr) _(—) ₁ *H _(corr) _(—)₂  [14]

where,M_(R) _(—) _(corr): correct reduced mass flow rate of the compressor 14;M_(R): reduced mass flow rate of the compressor 14;H_(corr) _(—) ₁: first corrective enthalpy value calculated with formula[13]; andH_(corr) _(—) ₂: second corrective enthalpy value.

In this case as well, it is important to stress that the reduced massflow rate M_(R) of the compressor 14 corresponds to the total air massflow rate m (preferably low) through the internal combustion engine 1used in the foregoing description regarding formulae [1] to [13].

The second corrective enthalpy value H_(corr) _(—) ₂ may be expressed asfollows:

H _(corr) _(—) ₂ =f(η_(scav) ,T _(air) /T _(ext))  [bypass duct 15]

where,H_(corr) _(—) ₂: second corrective enthalpy value;T_(ext): temperature of exhaust gases reaching the exhaust system andestimated by the electronic control unit;T_(air): temperature of air inside the intake manifold measured by atemperature sensor arranged inside the intake manifold; andη_(scav): scavenging efficiency of the cylinders 3 obtained from theratio between the total air mass flow rate m (preferably low) throughthe internal combustion engine 1 and the mass flow rate m_(com) of airtrapped inside the cylinders 3 that participates in combustion.

The second corrective enthalpy value H_(corr) _(—) ₂ is represented by amap based on the above-listed parameters that may be calibrated in apreliminary setup and tuning stage and be stored in the electroniccontrol unit.

According to a further variant, the correct reduced mass flow rate M_(R)_(—) _(corr) of the compressor 14 may be calculated as follows:

M _(R) _(—) _(corr) =M _(R)/η_(scav) *H _(corr) _(—) ₂  [16]

where,M_(R) _(—) _(corr) correct reduced mass flow rate of the compressor 14;M_(R): reduced mass flow rate of the compressor 14; andH_(corr) _(—) ₂: second corrective enthalpy value.

In this case as well, it is important to stress that the reduced massflow rate M_(R) of the compressor 14 corresponds to the total air massflow rate m (preferably low) through the internal combustion engine 1used in the foregoing description regarding formulae [1] to [13].

The second corrective enthalpy value H_(corr) _(—) ₂ may be expressed asindicated in formula [15], in which the second corrective enthalpy valueH_(corr) _(—) ₂ is represented by a map based on the above-listedparameters that may be calibrated in a preliminary setup and tuningstage and be stored in the electronic control unit.

It is also important to stress that the first corrective enthalpy valueH_(corr) _(—) ₁ and the second corrective enthalpy value H_(corr) _(—) ₂are completely independent from one another; or rather, the firstcorrective enthalpy value H_(corr) _(—) ₁ may be used to calculate thecorrect reduced mass flow rate M_(R) _(—) _(corr) of the compressor 14independently of the second corrective enthalpy value H_(corr) _(—) ₂,and vice versa.

In conclusion, the correct reduced mass flow rate M_(R) _(—) _(corr) ofthe compressor 14 may be expressed as follows:

M _(R) _(—) _(corr) =M _(R) *H _(corr)  [17]

H _(corr) =H _(corr) _(—) ₁ *H _(corr) _(—) ₂*(1/η_(scav))  [18]

where,M_(R) _(—) _(corr) correct reduced mass flow rate of the compressor 14;M_(R): reduced mass flow rate of the compressor 14;H_(corr): overall corrective enthalpy value;H_(corr) _(—) ₁: first corrective enthalpy value (which may be equal to1);H_(corr) _(—) ₂: second corrective enthalpy value (which may be equal to1); andη_(scav): scavenging efficiency of the cylinders 3 obtained from theratio between the total air mass flow rate m (preferably low) throughthe internal combustion engine 1 and the mass flow rate m_(com) of airtrapped inside the cylinders 3 that participates in combustion.

It is evident that the hitherto described method for correcting thereduced mass flow rate M_(R) of the compressor 14 may find advantageousapplication not only in controlling the solenoid control valve 35 of thewastegate 16, i.e. determining an objective position WG_(obj) of thesolenoid control valve 35 of the wastegate 16, but also in the case ofmechanical regulation of the wastegate 16 (for example, by means of anelectromechanical actuator) to determine an objective position WG_(obj)of the wastegate 16.

The above-described method for correcting the reduced mass flow rateM_(R) of the compressor 14 has numerous advantages.

Firstly, the above-described method for correcting the reduced mass flowrate M_(R) of the compressor 14 is simple and inexpensive to implementin an electronic control unit 21 of an internal combustion engine 1 asit uses measurements supplied by sensors that are always present inmodern internal combustion engines 1 and requires neither highcalculation capacity, nor large memory occupation.

Furthermore, the above-described method for correcting the reduced massflow rate M_(R) of the compressor 14 enables making control of thewastegate 16 robust, prompt and oscillation-free in all operatingconditions; in particular, even in operating condition where there is asignificant flow of air directly from the intake manifold to the exhaustof the internal combustion engine 1.

1. A method for correcting the reduced mass flow rate of a compressor(14) in an internal combustion engine (1) turbocharged by means of aturbocharger (12) provided with a turbine (13) and with a compressor(14); the internal combustion engine (1) also comprising an intakemanifold (4), an exhaust manifold (5) and a number of cylinders (2),each of which is connected to the intake manifold (4) by at least onerespective inlet valve and to the exhaust manifold (5) by at least onerespective exhaust valve; and in which the direct passage of air fromthe intake manifold (4) to the exhaust manifold (5) is allowed; themethod comprises the steps of: determining a control law (CL) thatprovides an objective opening of a wastegate (16) as a function of anactual supercharging pressure (P) and of a reduced mass flow rate(M_(R)) of the compressor (14); correcting the reduced mass flow rate(M_(R)) of the compressor (14) as a function of the enthalpy of a gasmixture flowing through the turbine (13) of the turbocharger (12) andcomprising both exhaust gases leaving the cylinders (2) and the freshair passing directly from the intake manifold (4) to the exhaustmanifold (5); determining an objective supercharging pressure (P_(obj));measuring an actual supercharging pressure (P); determining a firstopen-loop contribution (WG_(OL)) of an objective position (WG_(obj)) ofthe wastegate (16) by means of the control law (CL) and as a function ofthe objective supercharging pressure (P_(obj)); determining a secondclosed-loop contribution (WG_(CL1)) of the objective position (WG_(obj))of the wastegate (16) by means of the control law (CL) and as a functionof an actual supercharging pressure (P); and calculating the objectiveposition (WG_(obj)) of the wastegate (16) by adding the first open-loopcontribution (WG_(OL)) and the second closed-loop contribution(WG_(CL1)); and controlling the wastegate (16) so as to pursue theobjective position (WG_(obj)) of the wastegate (16).
 2. A correctionmethod according to claim 1, wherein correcting the reduced mass flowrate (M_(R)) of the compressor (14) as a function of the enthalpy of thegas mixture flowing through the turbine (13) of the turbocharger (12) iscarried out by means of the following formula:M _(R) _(—) _(corr) =M _(R)/η_(scav) *H _(corr) _(—) ₁ where M_(R) _(—)_(corr) correct reduced mass flow rate of the compressor (14); M_(R):reduced mass flow rate of the compressor (14); η_(scav): efficiency; andH_(corr) _(—) ₁: first corrective enthalpy value calculated by means ofthe following formula:H _(corr) _(—) ₁=[1+(T _(air) /T _(ext))*η_(scav)−1)] where T_(ext):temperature of exhaust gases reaching the exhaust system; and T_(air):temperature of the air measured inside the intake manifold (4).
 3. Acorrection method according to claim 1, wherein correcting the reducedmass flow rate (M_(R)) of the compressor (14) as a function of theenthalpy of the gas mixture flowing through the turbine (13) of theturbocharger (12) is carried out by means of the following formula:M _(R) _(—) _(corr) =M _(R)/η_(scav) *H _(corr) _(—) ₂ where M_(R) _(—)_(corr) correct reduced mass flow rate of the compressor (14); M_(R):reduced mass flow rate of the compressor (14); η_(scav) efficiency; andH_(corr) _(—) ₂: second corrective enthalpy value.
 4. A correctionmethod according to claim 1, wherein correcting the reduced mass flowrate (M_(R)) of the compressor (14) as a function of the enthalpy of thegas mixture flowing through the turbine (13) of the turbocharger (12) iscarried out by means of the following formula:M _(R) _(—) _(corr) =M _(R)/η_(scav) *H _(corr) _(—) ₁ *H _(corr) _(—) ₂where M_(R) _(—) _(corr) correct reduced mass flow rate of thecompressor (14); M_(R): reduced mass flow rate of the compressor (14);H_(corr) _(—) ₂: second corrective enthalpy value; η_(scav): efficiency;and H_(corr) _(—) ₁: first corrective enthalpy value calculated by meansof the following formula:H _(corr) _(—) ₁=[1+(T _(air) /T _(ext))*η_(scav)−1)] where T_(ext):temperature of exhaust gases reaching the exhaust system; and T_(air):temperature of the air measured inside the intake manifold (4).
 5. Acorrection method according to claim 3, wherein the second correctiveenthalpy (H_(corr) _(—) ₂) is calculated by means of an adjustable map,which is a function of the following parameters:H _(corr) _(—) ₂ =f(η_(scav) ,T _(air) /T _(ext)) T_(ext): temperatureof exhaust gases reaching the exhaust system; T_(air): temperature ofthe air measured inside the intake manifold; and η_(scav): efficiency.6. A correction method according to claim 2, wherein the efficiency(η_(scav)) is obtained from the ratio between the reduced mass flow rate(M_(R)) of the compressor (14), which represents the total air flowingthrough the internal combustion engine (1), and a mass flow rate(m_(com)) of air trapped in the cylinders (3) that participates incombustion.
 7. A correction method according to claim 6, wherein thereduced mass flow rate (M_(R)) of the compressor (14), which representsthe total air flowing through the internal combustion engine (1), iscalculated as follows:M _(R) =M _(com) +m _(scav) where M_(R): reduced mass flow rate of thecompressor (14), which represents the total air flowing through theinternal combustion engine (1); m_(com): mass flow rate of air trappedin the cylinders (3) that participates in combustion; and m_(scav) massflow rate of air that is not involved in combustion and flows directlyfrom the intake manifold (4) to the exhaust manifold (5).
 8. Acorrection method according to claim 1, wherein the enthalpy of the gasmixture flowing through the turbine (13) of the turbocharger (12) isobtained by means of the following formula:H(T _(mixt))=M _(R) *Cp*T _(mixt) where H: enthalpy of the reduced massflow rate (M_(R)) of the compressor (14), which represents the total airflowing through the internal combustion engine (1); M_(R): reduced massflow rate of the compressor (14), which represents the total air flowingthrough the internal combustion engine (1); Cp: specific heat of gas ata constant pressure; and T_(mixt): temperature of the gas mixtureflowing through the turbine (13) of the turbocharger (12).
 9. Acorrection method according to claim 8, wherein the temperature(T_(mixt)) of the gas mixture flowing through the turbine (13) of theturbocharger (12) is calculated as follows:T _(mixt)=(T _(ext) *m _(com) +T _(air) *m _(scav))/M _(R) whereT_(mixt): temperature of the mixture of exhaust gases and fresh airflowing through the turbine (13) of the turbocharger (12); M_(R):reduced mass flow rate of the compressor (14), which represents thetotal air flowing through the internal combustion engine (1); T_(ext):temperature of exhaust gases reaching the exhaust system; T_(air):temperature of the air measured inside the intake manifold (4); m_(com):mass flow rate of air trapped in the cylinders (3) that participates incombustion; and m_(scav) mass flow rate of air that is not involved incombustion and flows directly from the intake manifold (4) to theexhaust manifold (5).
 10. A correction method according to claim 1,wherein the direct passage of air from the intake manifold (4) to theexhaust manifold (5) is achieved by means of overlapping, with opportunetiming, the inlet valves that connect each cylinder (3) to the intakemanifold (4) and the exhaust valves that connect each cylinder (3) tothe exhaust manifold (5).
 11. A correction method according to claim 1,wherein the direct passage of air from the intake manifold (4) to theexhaust manifold (5) is achieved by means of a temporary deactivation ofa part of the cylinders (2), which are controlled to allow the passageof air from the intake manifold (4) to the exhaust manifold (5).
 12. Acorrection method according to claim 11, wherein the cylinders (3) aredivided into a number of active cylinders (3) to be controlled, when inuse, for injection and combustion and a number of inactive cylinders (3)to be controlled, when in use, to draw in a quantity of air in order toallow the air to flow from the intake manifold (4) to the exhaustmanifold (5).